In a powertrain test stand, the component to be tested (hereinafter also called test specimen), in most cases, is not necessarily connected to the environment, as is the case at its future site. For example, a powertrain on the test stand is connected to a driving machine and one or several load machines or—general—driven machines via relatively rigid shafts instead of being in contact with the road via tires. On the test stand, this mostly results in weakly damped (and thus pronounced) resonance frequencies, which the test specimen will not encounter at its actual site. If these resonance frequencies are excited by one of the machines, the resulting vibrations may heavily influence the test result or even result in destroying the test specimen and/or the test stand. Therefore, measures for damping these resonance frequencies on the test stand are necessary. The known methods will be listed below.
Above all, the use of softer, more damped connecting shafts may reduce the resonance frequencies and additionally also dampen them to a higher extent. If in normal operation the resonance frequencies are no longer excited, the problem will be solved. The low-pass effect of the soft connecting shaft turns out to be disadvantageous. In addition, quite a significant amount of power can be transformed in the shaft by friction (the shaft gets hot and might get destroyed).
Apart from such passive damping, methods for the active damping have been known, in which an additional torque corresponding to a shaft damping is applied on the driven machine. For example the differential angular velocity of the shaft, i. e. the difference between the measured speeds on the driven machine and the driving machine can be used for this (cf. e. g. DE 38 08 524 C2). The disadvantage of using the two current speeds on the driven side and the driving side is that these speeds are possibly available only in a noise-infested and considerably delayed way, e. g. due to the bus transmission time and/or filtering of the signal. In the worst case, applying such a distorted damping torque may cause the powertrain test stand to become particularly instable.
The need for high accuracy and dynamics when applying the two speeds was already recognized in EP 1 333 268 A2. It is proposed to estimate the differential angular velocity with the help of the measured shaft torque: The measured shaft torque is differentiated, weighted with a correction factor and supplied to the torque target value of the driven machine as correction value. The differentiation of a measured quantity, however, has the disadvantage that the constantly present measurement noise is intensified considerably. Indeed, the differentiated torque might be filtered by a low-pass filter, however, this method will get instable in the case of higher resonance frequencies, thus highly limiting the practical use. Furthermore, in the case of large test stands including a play and many distributed masses (e. g. in the case of powertrains with matching gears) only the first shaft close to the machine will be damped; the other shafts will remain undamped.
In addition, WO 2013/126940 A1 shows a method in which the measured shaft torque is not differentiated but as such is applied to the torque target value of the driven machine either directly or after a low-pas filtering. This measure, which initially comes as a surprise, results in very good damping in particular in the case of high resonance frequencies. However, if large masses are present on the test stand or the natural frequency to be damped is small, the attainable damping effect will be small.
Other, comparatively more complicated methods for the active damping show alternative options of using a measured torque (e. g. DE 102 47 347 A and U.S. Pat. No. 4,468,958 A) or try to predict the resonance behavior of the test stand (e. g. AT 010 301 U2 and U.S. Pat. No. 8,006,548 B2).
In connection with a highly realistic simulation or imitation of the behavior of a vehicle on a roadway, it has also been known to control the load machines of a powertrain test stand, i. e. the driven machines, by using a tire model. Either the speed (cf. AT 508 031 B1) or the torque (cf. EP 1 037 030 B1) of the driven machines may be controlled accordingly. It is the object of both methods to load the output of the powertrain more realistically and not to dampen the vibrations and, in particular, any natural frequencies of the test specimen on the test stand. Such a realistic load will obviously be achieved if the tire model is used for controlling the output (but not the input).
The US 2012/0166154 A1 relates to a slip simulation on the driven machines of a powertrain, whereby by using a tire model the test specimen on the test stand is to be subjected to a load correctly, i. e. realistically. Here, too, the reason for using a tire model on the test stand is to replace a real tire. The damping of natural frequencies, which is desired for a correct simulation, is accomplished by suitably setting a proportional and differential proportion in the speed controller.
The DE 10 2010 049 689 A1 also relates to a more realistic load of the test specimen if a real tire is replaced by a tire model. Its purpose is to simulate the tire by a better tire model on the test stand, whereby in particular by a detailed tire model, for instance, the flexing resistance, the influence of the tire temperature and the tire pressure, the proportion of roadway unevenness and dynamic tire loads etc. may be examined. Obviously, this makes only sense if instead of a real tire the tire model is used. In addition, the DE 10 2010 049 689 A1 indicates to determine individual controller parameters such as e. g. the P, I or D proportion of a PID controller on the basis of a tire model, a specific procedure not being disclosed here.
In the prior art, a tire model is used always expediently and with the sole purpose of replacing a real tire.
It is the object of the present invention to provide a method of the above given type or a device for carrying out this method, which avoids the above described vibrational problems in a simple and efficient manner or at least reduces them to an admissible degree. Problems in connection with time-delayed and time-shifted measured values are to be avoided and damping of relatively low-frequency vibrations is to succeed as well. In addition, the control is to be immune to measurement noise and achieve the desired damping also in transient examinations—in particular in the first examination of a test specimen.
To solve the object posed, the invention provides a method of the above given type, which is characterized in that the reference variable for the control of the torque of the driving machine is determined from a tire model of a virtual tire. In the control device indicated above, the invention accordingly provides that the outer control loop for controlling the reference variable of the inner control loop comprises a tire model of a virtual tire depending on a current speed of the driving machine. The invention is based on an unconventional and initially surprising use of the tire model in a place where normally (in a real setting) there are no tires. Therefore, the tire model is not used to replace a real tire (because such a tire does not exist in this place), but exclusively for its damping effect. The tire model used for controlling the driving, i. e. the driving machine of the powertrain test stand, does not result in a more realistic loading of the powertrain; only the damping effect of such a control is advantageously (also) used on the driving side. In this connection, the current speed of the driving machine (on which the virtual tire is positioned)—or an equivalent measured quantity—is inherently included in the tire model, however, it is not necessary to measure several speeds on the powertrain, so that problems resulting from the relative chronological behavior of several measured values may be avoided.
It is particularly favorable if the tire model establishes a preferably static correlation between the reference variable and a slip of the virtual tire or has such a correlation. The slip of the tire, which constitutes a prerequisite of an energy transmission, and a corresponding slip model are especially suited for the present method. In the case of a static correlation between the slip and the reference variable, i. e. the (modified) torque target value, influences of other—dynamic—measured quantities can be avoided completely.
In a particularly simple and thus preferred tire model, the reference variable is essentially calculated according to the formula MX=Fz·rdyn·D·sin(C·arctan(B·s)) and/or the outer control loop for determining the reference variable is essentially adapted according to this formula, with Fz being a contact force, rdyn a rolling radius, B, C and D constant tire-parameters and s a slip of the virtual tire. In this model, the reference variable shows a linear behavior only in a small area and the amount of abrupt changes is limited.
If a virtual roadway speed, which is included in the tire model, in particular in the slip of the virtual tire, is determined from an inverse tire model or if the outer control loop is adapted to determine a virtual roadway speed from an inverse tire model, the control of the driving machine (still) may be performed advantageously by predefining a desired torque. The inverse tire model can be chosen independently of the tire model. In particular, the inverse tire model does not have to correspond to the mathematical inverse of the tire model at all, but can be derived, for example, from a simplified tire model.
In this connection, it is advantageous if the virtual roadway speed is determined from the predefined value for the reference variable and a target speed of the driving machine, which target speed is preferably proportional to a target speed of the driven machine. Accordingly, the outer control loop is preferably adapted for determining the virtual roadway speed from a predefined value for the reference variable and a target speed of the driving machine. The exclusive use of predefined static quantities can prevent any feedbacks of dynamic measures quantities—besides the current speed of the driving machine—in the tire model.
Derivation of the inverse tire model is particularly simple if the inverse tire model makes or has a linear correlation between the virtual roadway speed and the target speed of the driving machine. The linear correlation can be used here, without having to incur any disadvantages with respect to the damping effect of the tire model, since the tire model (contrary to the inverse tire model) may very well have an at least partially non-linear correlation.